Rotary pump



Nov. 22, 1955 D, w, TRYHORN 2,724,340

ROTARY PUMP Filed May 2, 1950 5 Sheets-Sheet l 1955 D. w. TRYHORN 2,724,340

ROTARY PUMP Filed May 2, 1950 5 Sheets-Sheet 2 7 6 (3 5 fie//Ve/ "y W- 1 (3 4 Almaspkre by m Juana;

Nov. 22, 1955 D. w. TRYHORN 2,724,340

ROTARY PUMP F l May 2. 1950 5 Sheets-Sheet 3 f a 5 iii [y]! I E 5 1.9T if g .EzQeaZat: Mgumg 1111.027! k D. w. TRYHORN 2,724,340

Nov. 22, 1955 ROTARY PUMP Filed May 2, 1950 5 Sheets-Sheet 4 I'll Nov. 22, 1955 Filed May 2, 1950 5 Sheets-Sheet 5 I l o :2

United States Patent Ofihce 2,724,340 Patented Nov. 22, 1955 2,724,340 ROTARY PUMP Donald Wilfred Tryhorn, Slough, England, assignor, hy mesne assignments, to The British Internal Combustion Engine Research Association, Slough, England This invention relates to rotary pumps or motors of the lobe type.

The object of the present invention is to provide a rotary pump or motor of the kind referred to which will operate efficiently at high as Well as low pressures, be quiet in operation, will be less expensive to manufacture and will overcome the other disadvantages of the previously proposed devices.

The invention consists in a rotary pump or motor of the lobe kind having more than one member mounted for rotation in a casing, each or at least two of which rotate in opposite directions and in which each of the rotary members has lobe means, characterised in that at least one of the rotary members contains at least one port, one valve and one passage.

By way of example on internal combustion engines a more eflicient compressor would give more power for a given size and price of engine, a better fuel consumption and cooler engine running because of the lower inlet temperature. Compared with the centrifugal blower this compressor has the advantage of positive displacement of air for starting, the elimination of high speed gearing, and ease of matching to suit the engine requirements.

In its simple form a pump of the kind referred to has one lobe on each rotor but two or more lobes may be fitted. The advantages permitted by the new design include:

(1) Compression by action of the lobes while provid: ing adequate port areas at every point of the cycle,

(2) The reduction of the carry over of fluid from the clearance volume to the fresh charge and (3) The reduction of the leakage of compressed fluid past the lobes to the fresh charge.

Advantage (1) is achieved by reducing the ratio of lobe height to rotor diameter and then arranging the ports within the rotors themselves. Conditions (2) and (3) are satisfied by modifications to the shape of the lobes, or the use of two or more lobes per rotor.

In the following description of the invention reference will be made to its application to an air compressor, for the sake of clarity, but it is to be understood that a pump of this kind may be used as a displacing, compressing, or expanding machine by'the adjustment of i the period of operation of valves inserted within the rotors and controlling the port slots. It can also be adapted for use with other gases, vapours or fluids by changing the valve timing and the like as necessary.

The accompanying diagrammatic drawings illustrate by way of example only a number of embodiments of the invention in which Figures 1 to 4 are cross-sections of a pump showing four stages in the cycle of operation.

Figure 5 is a P. V. diagram of the pump of Figures 1. to 4. t

Figure 6 is a cross-section showing details of lobes.

Figure 7 is a section through a lobe radial of a rotor.

Figures 8 to 11 show various arrangements of rotors, bearings and passages.

Figures 12 and 13 show details of valves.

Figure 14 shows a construction having two rotors each with two lobes.

Figures 15 and 16 show constructions having three rotors.

Figure 17 shows a construction having unequal sized rotors.

Figure 18 is a representation of means to adjust a movable valve.

Figure 19 is a representation of means to automatically adjust the valves, and

Figure 19a is a representation of means responsive to atmospheric pressure for operating the inlet valve, and

Figures 20 and 21 show arrangements having a port in the casing.

Figures 1 to 4 show the lay-out diagrammatically the cycle of operation of a pump having two rotors each with a single lobe. The casing 1 is of figure eight form similar to that of a Roots blower. The rotors u, b take the form of two hollow cylinders running within this casing and geared to run at the same speed but in opposite directions by a pair of external gears not shown. (On each cylinder is a lobe 2, 3 which just clears the bore of the casing. The cylinders have slots 4, 5 cut in them on the leading side of the lobe of the delivery rotor b and the following side of the inlet rotor a each slightly greater in length than the lobes so that these may pass at each revolution. The slots 4, 5 break through to the centre bores 6, '7 of the cylinders and are used as ports for the supply and delivery of air. Ducts in one or both end plates of the outer casing are arranged to connect with the inner bores of the rotors. By making the bore of the delivery rotor circular and concentric with the rotor a stationary part tubular valve 3 can be inserted and the are covered by this valve determines the degree of compression obtained, before delivery port opening. Alternatively an automatic valve may be employed which opens when the pressure of compression is slightly higher than that in the delivery passage.

The suitability of this type of porting can be judged by a description of the cycle of operation with reference to Figures 1 to 4 which show the four distinct phases in the cycle. Figure 1 shows the commencement of the compression cycle. At this instant the whole casing is full of air at atmospheric pressure and the lobes are just coming within sealing distance of each other (i. e. a few thousandths of an inch clearance). Any rotation beyond this point reduces the volume of the casing on the leading side of the lobes and so that air is compressed. The escape of air from the delivery bore 7 is prevented by the tubular valve 8. Just as the volume in front of the lobes decreases so the volume behind them increases and in doing so a fresh charge is drawn in through the inlet port 6.

Compression continues until the required pressure is reached when, as shown in Figure 2, the delivery port 5 opens; the compressed air is then delivered until the point shown by Figure 3 is reached. From this point onwards, completing the cycle back to the point :as Figure 1, there is no change in volume so the delivery port 5 is arranged to close, as shown in Figure 3, and the compression cycle is completed.

On the inlet side however, the port and pipe losses cause an inlet depression and this is made good by leaving the inlet open during this dead period. This desirable feature is of double value in that it allows the inlet port to be kept open for the whole cycle, and thus no inlet valve is necessary, and no losses due to valve opening and closing times incurred.

The cycle will now be explained with reference to the P. V. diagram as shown in Figure 5. At point 1 the inlet commences and the rate of intake increases until point 2 the tip of the other lobe.

is reached; the effective lobe area being increased by the delivery lobe uncovering the inlet lobe. From point 2 the inlet is at a constant rate since the rotors turn at constant speed, therefore the inlet depression is constant. At point 3 the lobes come within sealing distance and as the effective lobe area is reduced, so the inlet depression decreases. Suction completed, the inlet depression is made good during the dead period, and may result in a pressure above atmospheric at the beginning of compression by reason of the centrifugal charging effect of the inlet port and wave actions in the inlet duct. At point 4 one complete revolution has taken place and the air is now on the compression side of the lobes. Compression takes place to point 5 where the delivery port opens, and the rate of opening being great, the pressure only rises a small amount between points 5 and 6 before settling down to the constant positive pressure required to move the air through the exhaust port. crease in effective lobe area reduces this velocity until, when it becomes zero, the delivery port closes. From point 8 to 9 the small quantity of air in the clearance space between the inlet lobe and delivery valve is discharged into the inlet air.

volume is carried round from cycle to cycle its effect is negligible.

A feature which cannot be shown on the P. V. diagram but which can be seen in Figure l, is that compression commences before the inlet port has moved out of the compression space. This is made possible by the speed of operation, since, by the time the increase in pressure becomes effective at the inlet port, the latter has passed the rotor centre lines, and the compression space is sealed. For example, if the lobe speed is one tenth of the speed of sound, then the rotor would turn one tenth of a rev. or 36, while the compression pressure wave is making one complete revolution of the rotor.

Any decrease in the clearance volume, in the delivery port at the end of compression, increases the efliciency of the cycle since it reduces the volume of hot com pressed air discharged into the inlet air. This volume may be decreased to a very small value if the lobe and port are shaped in a manner to be described. In its simplest profile the lobe is formed by an are as its outer edge while its concave side is formed by the path of The convexface of the inlet lobe is rounded to clear the edge of the delivery port. The inlet port is larger than the path of the delivery lobe, and the face of the exhaust lobe can be shaped to suit other considerations, i. e. similar to the inlet lobe for economy, shaped to give an easy air inlet during the dead period, or a more convenient distribution of masses for balancing the rotor. This design is satisfactory for low pressure ratios but at high pressuresit is preferable to reduce the clearance volume. Well rounded edges to the delivery port as shown at 9 and 10 in Figure 6 increases the orifice coefficient of the port and therefore allows the port itself to be smaller than a straight sides sharp edged one. To follow the curved edge 9, a wider based lobe is required, since the angles a. subtended at the centres of the lobes, should be the same. If in a particular design it is found that the crest of the lobe is an appreciable distance from the tubular valve then a further reduction in clearance space is'possible by increasing the outer diameter of the lobe, as shown in Figure 6 from 11 to 12. This necessitates a larger bore of casing on the inlet side and the removal of the tip of the lobe back to 13 to clear the delivery lobe. The leading corner 14 of this lobe is radiused to clear the tubular valve. By this means the clearance volume can be reduced to less than ,5 of the swept volume while still maintaining sutficient port area. A further advantage is that the larger lobe on the inlet rotor assists in balancing the large inlet port. When more than one lobe per rotor is used, the air in the clear- From point 7 to 8 the de- The volume from 9 to 1, is the clearance volume in the inlet port, and since this ance volume is discharged into the carry round volume before compression commences in that space.

The leakage of air from the compression space past the edges of the lobes to the inlet space reduces the cycle efiiciency, but these losses are greatly reduced if an alternative leakage path to the atmosphere is provided. This condition is achieved by cutting a slot 15 round each lobe communicating with ports in the end walls of the casing. These ports can also be arranged to carry away air which would otherwise leak past the ends of the rotors, and are shown in Figure 7. A preferred form of passage comprises a hole 16 beneath the surface of the lobe, which is connected with the surface by a narrow slit 15. This slit is sufficient to pass the leakage air, but not large enough to allow any large escape of air during the inter lobe sealing period. The hole at the bottom of the slit is of sufiicient area to pass the leakage air without excessive back pressure.

Since it is the temperature of the leakage air which lowers the cycle efliciency an alternative is to pump cold compressed air, or a fluid, out through theslits, thus closing the gap to the lower pressure air in the compression process. Some cycle cooling can also be obtained with this method. When more than one lobe on the rotor is used the leakage passes to the carry round volume and is not lost. 1

There are a number of methods which can be used for supporting and locating the rotors and three of these will be mentioned briefly. The first and simplest form, as shown in Figure 8, is to overhang the rotors. This gives a very simple construction with an absence of ducts but necessitates the use of short rotors and a wide spae ing between the bearings. Improved rotor support can be obtained by continuing the shaft through the passages to an outer bearing as shown in Figure 9 in which case end ducts are required. To obtain a continuous air delivery a pair of rotors may be put back to back, as shown in Figure 10 and the lobes set apart. With this arrangement a continuous air delivery can be achieved at pressures of the order of 25 lb. sq. in., this is accompanied by a smoother driving torque and smaller bearing loads. Holes joining the discharge passage balance the flow through the end ducts so reducing the discharge losses. As a further alternative a pair of rotors may be overhung as shown in Fig. 11. a

In the cases shown in Figures 9 and 10 the ducts may be of substantially volute diffuser type, collecting or delivering the air with the same direction of rotation as the rotors. Pre swirl or straightening vanes can be fitted in the mouth of the inlet and delivery bores to improve the air flow.

The tubular delivery valve 8 is supported by the main casing and a small clearance is maintained between the valve and the bore of the rotor. It may take the form of a smooth arc of a tube or have serrations 17 machined in its outer edge to improve air sealing as shown in Figure 12.

The length of arc of the delivery valve may be controlled if part of it is made in the form of a grid 18 behind which a second tubular type valve 19 can be rotated, see Figure 13. The position of this inner valve would then determine the point at which the delivery port opened, and may be controlled manually or automatically to ensure that the compressor works at the optimum efficiency for the delivery pressure existing at any instant. Further reference to such a construction will be made later.

Figure 14 shows the condition which arises when two lobes per rotor are fitted; in space X delivery of compressed air is taking place, in Y a fresh charge is being taken in, and in the spaces Z the previous fresh charge is trapped and is being carried round between the lobes. Underv these conditions the higher pressure air leaking past the lobes will raise the pressure of this air before its compression cyclecomrnences. Similarly heat tive displacement compressors.

about decreasing the time x area of the delivery port.

transfer from the metal surfaces will raise the pressure and, in taking in this heat, the charge cools these surfaces, thus reducing distortion. By the use of this arrangement the total volume of the inlet zones is equivalent to four of the spaces Z per revolution, and is greater than can be obtained with the single lobe type.

The most important loss in compressors which maintain a clearance space between all Working parts is leakage-in this case leakage past the lobes and between the drums. Because of the rolling action the clearance between the drums may be very small, but a reasonable clearance has to be maintained between the lobes and the casing. The great advantage of the two lobe design is that this leakage is made to perform a first stage of compression. During the majority of the pressure part of the cycle the leakage past the lobes enters'the carry round volume of the nextcycle thus raising its pressure. The trapping of this leakage air is doubly advantageous in that it prevents mixing with the inlet-air. Should mixing take place the inlet-air would be raised in temperature, and thus the work of compression and the delivery temperature would also be raised.

A second feature of the carry round period is that the fresh charge is brought into contact with the metal walls heated during the previous compression and delivery periods. This has the effect of reducing thermal distortion which is often a serious problem in posi- The effect of this heat exchange on the cycle efficiency is negligible. The gain of pressure by pre-heating the charge at constant volume and the advantage of cooling the compression process are counteracted by the increased work required to compress the charge from the higher starting temperature.

The volumetric displacement of the two lobe design is equal to four of the carry round volumes per revo1ution minus the inter lobe clearance space, and is greater than. the complete annulus of the single lobe type by To take full advantage of this increase it is advantageous to cut ports through the corner of the casing to by-pass the lobes in the position shown in Figure 19. This allows the right hand volume to fill with air as the tip of the inlet rotor lobe withdraws from this space.

Similar lobe by-pass ports may be cut through the top corner of the casing to balancethe small pressure difference occurring at the beginning of direct compression. The value of these ports becomes less as the delivery pressure is increased since the delivery valve is not then opened until well after the joining of the two zones and the slight unevenness at the beginning of compression has a negligible effect on the overall efliciency. The bypassing of the two sharp corners by ports improves the air flow within the casing thereby reducing the inlet and delivery losses.

A valve is necessary in the inlet rotor to seal the compression zone during the early part of compression and is extended to about 180 of arc to seal the carry round volume and so allow the build of leakage compression. Since the inlet process only takes half a revolution this valve does not cause any appreciable reduction in the port time x area value.

The arc covered by the delivery valve is much longer than for the single lobe type and since the rate of port opening and closing are the same, this has the effect of The reduction is not great for low pressure settings but may limit the maximum pressure obtainable to less than that for the single lobe type.

Two lobes and ports in diametrically opposite positions .form an intrinsically balanced system. This is advantageous as it gives symmetrical stressing and distortion.

The fact that balancing holes are not required allows greater freedom in design and larger lobes, ports and rotor bore can be accommodated. The two cycle p'er revolution greatly reduce the variations in driving torque with the 6 result that no rotor flywheel effect is required arid light alloys may be used. At low pressures the air delivery is almost continuous and is delivered at constant velocity for the majority of the cycle.

Three two lobe rotors in a line give four cycles per revolution compared with only two for the two rotor type, but suffer the disadvantage of starting the compression late in the rotation of the rotors because of the incomplete encasing of the centre rotor. In this case the compression does not begin when each lobe tip dips into a pocket of fresh charge but when the lobe tip passes the intersecting point of the two bores in the casing. Apart from the increased displacement this type has the ad vantage that the centre rotor is balanced for pressure loads on the bearings. The bore of this rotor is used as either a common delivery port or as a common inlet port depending on direction of rotation. In this latter case it is advantageous to omit the inlet valve and increase the bore over that of the delivery rotor to reduce the losses incurred by passing the larger volume of air.

The are of rotation required for a compression and delivery cycle becomes less as the number of lobes is increased and if three lobes per rotor are fitted the three rotor machine can be made to give a cycle almost similar to that of the two lobe rotormachine (2 x 2) apart from the reduction of leakage compression periods. Such a machine is highly competitive with the 2 x 2, the greater displacement per revolution compensating for the reduced speed which is necessary for the same port time area. Either outer or inner rotors may contain the delivery passage depending upon the direction of rotation as shown in Figures 15 and 16. An advantageous feature of that of Figure 16 is that the single delivery passage is capable of passing the alternating delivery periods with very little interference.

More than three lobes per rotor may be used but is accompanied by reduced port time areas. The various rotors may be fitted with different numbers of lobes providing the timing gears have the numbers of teeth in proportion to the number of lobes on the associated rotor.

The limitation on the addition of another rotor to a compressor of this family is that the interference of the casing bores should not appreciably reduce the arc of casing existing and required to maintain the seal of the space in which the compression and delivery take place. This are becomes progressively smaller as the number of lobes per rotor is increased. It has been shown that this interference takes place if two lobes are used and a third rotor added. Three lobes give zones which just avoid interference if three rotors are put in a straight row. The length of the centre are of casing is increased if the centre rotor be of larger diameter than the outer rotor and another lobe arrangement suitable for three rotor machines is two on each outer rotor and three on the centre. This allows elfective port areas as large as the normal 2 x 2 machine.

Figure 17 shows such a construction having three rotors, the centre one having three lobes While the outer rotors have two lobes on each. The inlet port is provided in the centre rotor while delivery takes place via the ports in the outer rotors. Suitable gearing is provided to give appropriate speeds of rotation between the inner rotor and outer rotors.

Providing the minimum sealing are referred to is maintained there is no reason why the third or subsequent rotors should lie on the same centre line as the first two. The result of this is that if at least three lobes are used per rotor, rotors can be added until a complete ring is formed. Such a machine works providing an even number of rotors is used to maintain the right direction of rotation. Other geometrical arrangements can be made up as can be made up with sets of meshing gear wheels.

The reduced port areas given when more than two lobes per rotor are fitted make it increasingly advantageous to open and close the ports more rapidly. The

natural speed of this occurrence is the surface speed of the rotor bore but since it is the relative velocity of the port to the valve which determines the rate of opening this can be increased by rotating the tubular valve in a direction opposite to the rotation of the rotors, e. g. a known means of converting the steady rotary motion of the rotors to an oscillating motion could be used to rock the tubular valve; for example, Figure 18 shows a crank 20 driven from the rotor .via the gears 21 and 22 and shaft 23. The extent of the throw of the crank 20 is shown as being adjustable by the diagrammatically represented mechanism 24 while the phase relationship of the valve 25 and the rotor 26 is adjustable by means of the diagrammatically represented mechanism to charge the angular phase relationship of the gear Wheels. This crank mechanism may be geared to rotate once during the valve opening period so that the valve is rotating towards the approaching port at the port opening point then swings back during the port open period and be again approaching the port at the closing point. With such an arrangement variation of the phasing and of the crank radius can be used to control the amount of compression given before delivery commences.

By adjustment of the valves this machine may become a compressing, expanding or simple displacing machine and thus it may be used for any purpose requiring any of these processes. Some such uses are; superchargers for internal combustion engines, pressuriscr for aircraft cabins, compressors for gas turbine or gas generator systems, displacement pump for liquids, compressors for industrial purposes, exhaustors for vacuum equipment or as motors delivering power from the displacement or expansion of gases or vapours.

.The advantage of the proposed machine for any purpose involving compression expansion or displacement of a fluid, are its purely rotary motion, high efliciency and small size. Other advantages accrue for the particular uses quoted and these will now be stated in more detail. The working principles of the machine are the same for all these purposes and the differences lie only in the settings of the valves. Figure 19 shows a cross section of the machine through the rotor with adjustable tubular valves 27, 28 within each rotor. Rotation of the adjustable valve segment in the inlet rotor 29 in an anticlockwise direction reduces the length of the inlet period and so controls the quantity of fluid passed by the machine. Rotation of the adjustable valve segment within the delivery rotor 30 in an anticlockwise direction increases the degree of internal compression carried out before delivery commences. The delivery pressure is determined by the relative capacities of the compressor and supplied machine and for efiicient operation the internal compression should be approximately equal to this delivery pressure. To obtain this the rotatable segment of the delivery valve is coupled to a pressure sensitive device 31 operated by the difference between the delivery pressure and the compression pressure.

Internal combustion engines require supercharging only at high loads and because present superchargers give the same supercharging at all loads there is an unnecessary waste of power at light loads. This fault can be over come with the present design by connecting the valve 27 in the inlet rotor 29 to the engine throttle as shown by the arrow 32 so that the quantity of air delivered per cycle is reduced as the throttle is closed. On petrol engines the quantity of fuel supplied is proportional to the quantity of air delivered and in this case a wide range of power control is given and over most of the working range this control takes the place of the normal throttle valve. Such an arrangement is especially advantageous in that it eliminates the throttling loss and therefore increases the cycle efficiency at light loads.

In one particular case it may be arranged, by making the blower inlet volume less than the engine swept volume 1 at full, load, that the inlet pressure he still below atmospheric. In this case the compression ratio of the engine can be raised and the cycle efiiciency increased without'increase in maximum pressure. Heat from the exhaust gases or any other known method may be used to keep the fuel air mixture warm enough for combustion temperature to be reached in the cylinder after compression.

Because the two stroke diesel engine requires both a high supercharge and a large excess of scavenge air to allow it to operate at high loads, the supercharger power is high. At light loads it takes a large part of the total output and because of this the fuel consumption is high. By using a supercharger of the proposed design the inlet control valve may be coupled to the engine throttle so that a greatly reduced quantity of air be delivered to the engine at light loads; the supercharger then requires less power and the fuel consumption of the engine is improved.

For maximum efliciency these superchargers have the delivery port control valve operated by a pressure sensitive device in the supercharger delivery, but this may be omitted when the desired change in air quantity is small. This is because the reduction in air aspirated causes a lower pressure at the beginning of compression and so a lower delivery pressure, which is What is normally required for the reduced air flow in the system.

Alternatively the compressor may be used in the same way as normal positive displacement compressors by leaving the inlet or both valves fixed.

In addition to the advantage of supplying air free from oil contamination, the proposed compressor fulfils other requirements. To maintain the pressure in an aircraft cabin near atmospheric pressure requires a steady increasing compression as the altitude of the aircraft increases. At the same time the weight of air delivered to the passengers should remain constant. These conditions are obtained from the proposed design by rotating the controllable inlet port valve in a clockwise direction as the altitude increases, so that a constant mass of air is taken in per cycle. This may be achieved by connecting the valve by a linkage to a device (Figure 19a) operated by the pressure of the atmosphere around it.

Useful work may be done in discharging the spent air from the cabin to the atmosphere and if this is desired the proposed machine can be operated as an expanding machine being used as an air motor. This may be a separate unit or integral with the pressuriser.

The proposed compressor is especially suited to use with gas turbines and gas generators as it is capable of working at a high pressure ratio by reason of its internal compression and because of its freedom from surging. Operation of the control valve in the inlet port by a linkage to the fuel system allows rapid changes of load on the set without overheating the turbine blades.

By reducing the length of the delivery valve so that it is open for the length of the inlet suction period of the machine may de described as a displacing machine. In this condition-it is capable of acting as a pump for incompressible fluids and has the advantage over normal pumps of improved volumetric efliciency. This improvement is given by the centrifugal charging effect of the rotating inlet port which greatly reduces losses due to cavitation which is often experienced with the normal gear type pump.

The inlet control valve with associated delivery valve control may be used if variable throughput at constant speed is required.

The high operating speed, small size, and low cost of the proposed type of compressor makes it attractive for use as a normal industrial compressor for use with pneumatic tools or any process where air gas or vapour is required at a pressure.

Again if constant speed is essential the quantity of air delivered can be varied by the regulation of the movable control valves.

The high internal compression ratio which can be obtained makes this type of compressor suitable for use as an exhauster. In this case the machine takes in gas at a very low pressure, compresses it up to atmosphere pressure and then delivers it at atmosphere pressure. For normal systems the efficiency of the system during the creation of the vacuum is of little importance so fixed valve times can be used. The controllable valves can be used if other conditions are required such as control of mass flow or eflicient running with different degrees of vacuum. A refinement for this type of pump is the use of either the controllable inlet port timing, delivery port timing, or a blow-oil valve so that excessive pressures are not built up in the machine during the creation of the vacuum, when the inlet pressures would be near atmospheric.

Most compressors can be used as motors if a supply of gas or vapour under pressure is available, but only types giving internal expansion are capable of giving high efliciences. In the proposed machine internal expansion is obtained by closing the inlet valve early and allowing the working fluid to expand to the end of the normal inlet period. No internal compression is required so the delivery valve is arranged to be open for the whole of the period'during which the lobes are approaching each other, this then forms the exhaust stroke. If a controllable inlet valve be fitted then variable cut oil can be used as on steam engines and used for this purpose, the motor is anjintermediate machine between the steam engine and turbine.

Figure 20 shows a construction of a pump having two rotors each with two lobes in which a hole 33 is provided in the casing through which air is drawn from the atmosphere. In this embodiment the inlet port in the rotor is dispersed with and a single port provided in the casing. This rotor is shown solid, but is provided with indentations 34, 34 to allow of the passing of the delivery rotor lobes. This arrangement enables an extremely large port area to be used, the consequent reduction in suction loss may outweigh the loss of the centrifugal charging effect and the increased loss during the inter lober sealing period.

In the alternative arrangment of Figure 21 the inlet port 33 is in the casing while both rotors contain delivery ports 35, 35, and although the clearance volume is increased, thus loweringthe efficiency, the throughput is increased by reason of the higher rotor speed for a given gas velocity through the ports. The ports in one of the rotors b are arranged as previously described in advance of each lobe while the ports in the other rotors a are cut through each lobe on the cylinder immediately preceding each lobe, the delivery port in the rotor a being fitted with a control valve 36 to determine the point of opening as with the normal delivery rotor b. Rotor a is provided with indentations 36, 36 to accommodate the lobes of rotor b.

The scope of the invention is not limited to the various embodiments hereinbefore described with reference to the accompanying figures, but various modifications can be eliected without departing from the scope of the invention.

I claim: i t

1. In a rotary fluid machine, the combination of a casing forming a plurality of circumferentially-intersecting cylindrical chambers defining a common circumferential opening therebetween, companion rotor cylinders mounted in said chambers adapted to be rotated at the same peripheral speed in opposite directions, said cylinders being of smaller diameter than the chambers to provide a working space intermediate the cylinders and the chambers and said cylinders being in rolling contact with each other throughout a portion of their circumference, radially-projecting, cooperating lobes upon the cylinders having close-fitting clearance with the walls of their respective chambers, a fluid passage through one of said cylinders, a port adapted to receive a lobe on a companion cylinder extending between said fluid passage and the circumference of its cylinder to communicate with the working space lying between one side of the cooperating lobes of a pair of companion cylinders,

. 10 said port having a minimum circumferential width greater than the radial height of the cooperating lobe disposed above the periphery of said companion cylinder, and a second fluid passage communicating; with the working space lying between the other side of the cooperating lobes of said pair of cylinders.

2. A rotary fluid machine as set forth in claim 1 wherein the radial height of said cooperating lobe is such that its crest extends into the port of the companion cylinder to a position closely adjacent the fluid passage during rotation of the rotor cylinders.

3. A rotary fluid machine as set forth in claim 1 where in the port extending between the fluid passage and the circumference of its cylinder is located adjacent the lobe upon the same cylinder.

4. Rotary fluid machine as claimed in claim 1 in which at least two of the rotary members each have a longitudinal passage at least one of which serves for the entry of the working fluid into the machine while at least one other serves for the exit of the working fluid, at least one of the longitudinal passages being circular in cross section and coaxial with the rotary member and containing a valve of circular cross section which is cut away to form the required port and is mounted coaxially with the rotary member so as to co-operate with at least one communicating passage to control the passage of fluid therethrough, said valve being provided with a closely fitting concentric tubular member cut away to form the required port, the point of closing of the entry port being adjustable.

5. Rotary fluid machine as claimed in claim 1 in which at least two of the rotary members each have a longitudinal passage at least one of which serves for the entry of the working fluid into the machine while at least one other serves for the exit of the working fluid, at least one of the longitudinal passages being circular in cross section and coaxial with the rotary member and containing a valve of circular cross section which is cut away to form the required port and is mounted coaxially with the rotary member so as to co-operate with at least one communicating passage to control the passage of fluid therethrough, said valve being provided with a closely fitting concentric tubular member cut away to form the required port, the point of opening of the exit port being adjustable.

6. Rotary fluid machine as claimed in claim 1 in which at least one of the longitudinal passages is circular in cross section and is coaxial with the rotary member and contains a valve of circular cross section which is cut away to form the required port and is mounted coaxially with the rotary member so as to co-operate with at least one communicating passage to control the passage of fluid therethrough, said valve being provided with a closely fitting concentric tubular member cut away to form the required port, said tubular member being adjustable so as to position by rotation about its common axis with the rotary member, the adjustable tubular memher being automatically controlled by pressure sensitive means.

7. Rotary fluid machine as claimed in claim 1 in which at least one of the longitudinal passages is circular in cross section and is coaxial with the rotary member and contains a valve of circular cross section which is cut away to form the required port and is mounted coaxially with the rotary member so as to co-operate with at least one communicating passage to control the passage of fluid therethrough, said valve being provided with a closely fitting concentric tubular member cut away to form the required port, said tubular member being adjustable as to position by rotation about its common axis with the rotary member, the adjustable tubular member being rotatable cyclically relative to the rotation of the said rotary member.

8. Rotary fluid machine as claimed in claim 1 in which at least one of the longitudinal passages is circular in cross. section and. is coaxial with the rotary member and contains a valve of' circular cross section which is cut away to form the required port and is mounted coaxially with the. rotary member so as to co-operate With at least one communicating passage to control the passage of fluid therethrough, said, valve being provided with a closely fitting concentric tubular member cut away to form the required port, said tubular member being adjustable as to position by rotation about its common axis with the rotary member, the adjustable tubular member being rotatable cyclically relative to the rotation, of the said rotary member, and the phase relationship of the same and the rotary member being adjustable,

9. Rotary fluid machine as claimed in claim 1 in which at least one, of, the longitudinal passages is circular in cross section and is coaxial. with the rotary member and contains avalve. of circular cross section which is cut away to form the required port and is mounted coaxially with the rotary member so as to co-operate with at least, one communicating passage to control the passage. of fluid therethrough, said valve being provided with a. closely, fitting concentric tubular member cut away. to form the required port, said tubular member being adjustable. as to position by rotation about its common axis with the, rotary member, the adjustable tubular member being rotatable cyclically relative to the rotation of the. said. rotary member, and the extent of the arc of movement of. the same being adjustable.

10. Rotary fluid machine is claimed in claim 1 in which each rotary cylinder has at least two lobes and aport is providedin the casing of the machine.

11. Rotary fluid machine as claimed in claim 1 in which each rotary cylinder has at least two lobes and a port is provided in the casing of the machine, at least one fluid passage being for the exit of the fluid from the machine, entry being byway of the port in the casing.

12. Rotary fluid machine as claimed in claim 1 in whichthe rotary cylinders of two machines are provided on opposite ends of, common axles with one set of heatings and gears between them.

13. In a rotary fluid machine, the combination of a casing forming a plurality of circumferentially-intersecting, cylindrical chambers defining a common circumferential opening therebetween, companion rotor cylinders mounted in said. chambers adapted to be rotated at the same peripheral speed in opposite directions, said cylindersbeing of smaller diameter than the chambers to provide a working space intermediate the cylinders and the. chambers and said cylinders being in rolling contact with each other throughout a portion of their circumference, radially-projecting, cooperating lobes upon the cylinders having close-fitting clearance with the walls, of their respective chambers, each of said cylinders being provided. with a fluid passage and a port adapted to receive a lobe on a companion cylinder extending between. said fl'uid passage and the circumference of its cylinder, said ports being located upon opposite sides of said lobes considered in the direction of rotation of said cylinders, and said ports each having a minimum circumferential width greater than the radial height of the cooperating lobe disposed above the periphery of said companion cylinder.

14. A rotary fluid machine as set forth in claim 13 wherein one fluid passage and port provides means for admitting working fluid into the machine and the other fluid passage and port provides means for exhausting working fluid fromthe machine.

15'. In a rotary fluid machine, the combination with a casing forming a plurality of circumferentially-intersecting cylindrical chambers defining a common circumferential opening therebetween, companion rotor cylinders mounted in said chambers adapted to be rotated at the same peripheral speed in opposite directions, said cylinders being of smaller diameter than the chambers to provide a working space intermediate the cylinders and the chambers and said cylinders being in rolling contact with each other throughout a portion of their circumference, radially-projecting, cooperating lobes upon the cylinders having close-fitting clearance with the Walls of their respective chambers, eachof said cylinders being provided with a fluid passage and a port adapted to receive a lobe on a companion cylinder extending between said fluid passage and the circumference of its cylinder, said portsv being. located adjacent and upon opposite sides of said lobes. considered in the direction of rotation of said cylinders, and said ports each having a minimum circumferential Width greater than the radial height of the cooperating lobe disposed above the periphery of. said companion cylinder, and a valve for controlling the passage of fluid through one of said ports.

16. A rotary fiuidmachine as set forth in claim 15 wherein thefluidpassage of one of said rotor cylinders is concentric with. the, axis of rotation of said cylinder and, said valve comprises a. stationary, arcuate member having close. clearance engagement with the inner circumferenceof said. fluid passage for intermittently opening and closing the port in said cylinder during its rotation.

References Cited in the file of this patent UNITED STATES PATENTS 883,894 Knowles Apr. 7, 1908 1,076,299 Marshall Oct. 21, 1913 2,559,590 Brown July 10, 1951 FOREIGN PATENTS 22,259 Great Britain of 1904 55,885. Sweden Jan. 15, 1924 341,324 Great Britain Jan. 15, 1931 627,202. France.- May 30, 1927 665,484 Great Britain Jan. 23, 1952 

